Hydraulic snubbing devices for ameliorating the rolling motion of railway cars are commonly used. A common practice is to substitute the hydraulic damper for one of the springs of the spring group at each side of the car between the bolster and the side frame of the truck. The following United States Patents are illustrative of hydraulic snubbing devices of this character: U.S. Pat. Nos. 3,773,147, 3,837,292, 3,868,912 and 3,874,307.
The circumstances under which such devices must operate are quite severe. It is necessary that they have a long operational life with little opportunity for periodic service. The ambient conditions (dust, water, etc.) to which they are subjected are likely to have a damaging effect. A particularly significant problem is caused by the fact that the loads are not constantly applied axially of the snubber, but the transverse movements of the bolster with respect to the side frame, and vice versa, act to apply non-axial loads to the snubber. Thus, the snubber must be able to resist such forces that tend to cock the piston rod and piston in the members that define the cylinder and piston rod bearing and to withstand the added wear caused thereby. Furthermore, the heat that snubbers must generate and dissipate to the atmosphere results in elevated temperatures which reduce seal life. These temperatures can be of sufficient magnitude to preclude the use of certain seal materials which could yield optimum service life provided only that the operating temperature were lower. Last but not least of the significant considerations as to such snubbers is the matter of manufacturing cost. It is, of course, very important that the cost be as little as possible while still producing a device that will have the required performance, including service life, etc.
The principal object of the present invention is to provide a hydraulic snubber apparatus for railroad cars which will have excellent characteristics under the operational conditions discussed and yet which will have a relatively low manufacturing cost as compared to commercially available devices. A number of features of the present invention contribute toward this end.
In any hydraulic cylinder system, numerous machining operations are required to produce a structural configuration which will provide for the performance of various functions. These functions which will be treated specifically as regards the primary structure embodied in this invention are the bearing support of the piston rod, the bearing support of the piston, high pressure static sealing, low pressure dynamic sealing, and low pressure static sealing.
The mechanical fit at adjacent structural surfaces which accommodate these functions governs to some extent either or both the integrity and durability of the product as it depends upon these functions. This situation is in particular evidence at the piston rod seal where eccentric loading of the seal by the rod can be a major cause of accelerated seal wear. This eccentric loading of the rod seal results from a side load component on the rod which causes the rod to move laterally through excessive free play in the interface fit between the rod and the rod bearing. This free play increases with wear of the rod bearing which in turn is accelerated by free play between the piston bearing surface and the cylinder bore surface. Lateral free play at both of these reciprocating interfaces, rod and rod bearing as well as piston bearing and cylinder wall, tends to increase wear due to the higher contact stresses; and as a result, free play tends to be self-increasing.
The degree to which an optimum fit may be obtained between any fabricated functional interfaces is determined by machining tolerances, and therefore manufacturing costs, not only at the specific interface portion in question, but also at any other functional interfaces which mutually establish the geometric constraint of any member within the assemblage of all members. All fits must be such that the machine can be assembled without damage to components while assuring satisfactory performance of the various functions. This implies the evidence of minimum free play fits, minimum interference fits, or maximum interference fits, depending on the specific function involved.
In addition to considerations of the dynamic low pressure rod seal function discussed before, the fit between the reciprocating and stationary members must permit freedom from any jamming condition which cannot be overcome by the force of the return spring. Furthermore, the fit of the reciprocating members must be such that internal movements due to eccentric loading of the rod are not supported by either the rod bearing or the piston bearing alone, a situation which can cause either or both failure and premature wear of these components.
In addition to the dynamic functional interfaces discussed above, the performance of the static functional interfaces must also be included in overall fit considerations. In order to avoid degradation of hydraulic performance, there must be sufficient interface contact at the high pressure seal to provide adequate obturation of the head end of the cylinder bore. Low pressure static sealing interfaces on the other hand require close fits in order to ensure adequate compression of elastomer seals to compensate for partial memory loss due to operational and environmental thermal histories.
Machining tolerances may be separated into three general categories including feature size, feature form and feature relationships. The first category pertains to such dimensional aspects as feature lengths and bore diameters. The second involves feature characteristics such as flatness, straightness, and roundness. The third category includes considerations such as parallelism of surfaces, squareness of circular bores to planes and concentricity of circular bores with respect to one another, a major consideration in hydraulic cylinder design.
One aspect of this invention is its structural configuration which offers an economical realization of optimum fits at all functional interfaces. This is accomplished through a configuration which permits articulation of components such that there is a minimum of interdependency on multiple machining operations in establishing any functional interface fit. In this regard, principal attention should be focused on machining tolerances in the third category, specifically that of concentricity of bores. In addition, moreover, the configuration permits economic utilization of precision machining operations in establishing all significant dimensional and characteristic tolerances pertaining to the first and second categories in combination.
The cylinder body is a single unit into which is pressed the rod bearing. In a preferred embodiment the interface of the two members is an extension of the surface defining the cylinder bore, and therefore the accuracy of the positional relationship of the internally pressed surface and the cylinder wall is obtained in a single machining operation on an open-ended bore. As a consequence, the only machining operations which can introduce bore eccentricities limiting the allowable closeness of fit at interfaces between reciprocating and stationary members are confined to the internal bore of the rod bearing and the rod pilot in the piston. The high pressure end of the cylinder is closed by a head member which serves as the reaction member contacting the side frame of the car truck. This member is rigidly fastened to the cylinder at the bottom flange of that structure by virtue of a multiplicity of flange screws.
With the addition of the piston rod which is bolted to the piston in situ at assembly, all aforementioned components comprise the system primary structure and as such transmit all thrust loads, including any prevailing transverse component, between the car bolster and car truck side frame. It can be established that this arrangement of a single-ended cylinder primary structure provides for specification of the minimum acceptable clearance fit between reciprocating and stationary members for any given specification of machining tolerances. Furthermore, it will be seen in the following discussion that the primary structure configuration provides for an effective accommodation of secondary functional interfaces without compromise of the aforementioned fit. Also, with the exception of the press fit of the rod bearing, all fastening in the primary structure is accomplished through the use of relatively long screws with relatively short thread engagement lengths. This approach provides one of the better means of obtaining positive, vibration resistant structural fastening.
A novel approach is used to provide a dimensionally compact solution to high pressure static sealing and return valve seating without requiring excessive fabricating precision. These functions are accommodated by a single ring member located generally in the annular recess in the cylinder head member. The ring is manufactured with an outside diametric dimension such that there is a slight interference fit with the cylinder bore. This fit occurs at the edge juncture of the cylinder bore and the small chamfer at the head end of the bore and is accomplished when the reaction head is assembled to the lower cylinder flange by means of the flange screws. Prior to this operation the ring is free to "float" in the recess and therefore is self-locating with respect to the aforementioned edge. This approach to high pressure static sealing is thus free from most fit problems related to machining eccentricities and takes full advantage of the diametrical precision usually associated with a straight through cylinder bore honing operation. The edge or line contact sealing interface which permits the ring to have a low profile height is further relieved of manufacturing tolerance requirements by providing the ring with a self-sealing geometry under hydraulic pressure. The ring is of generally "U" or "L" shaped cross section, the base leg being somewhat lower than the sealing edge. As a result the exterior wall of the ring tends to deflect radially outwards under hydraulic pressure thus further obturating the cylinder bore. The top inner portion of the ring serves as the seat for the return flow check valve which is comprised of a thin elastically deformable plate located laterally by a lug protruding through a central hole in the plate, such lug being a portion of or a separate member joined to the reaction head member. The low profile height of the ring, by virtue of its minimal intrusion into the working chamber space of the cylinder bore, offers a maximal piston skirt length and therefore a maximal spread between the rod bearing and piston bearing contact points. This feature is important in minimizing bearing stresses, and therefore bearing wear, since it provides for both the reduction of bearing loads as well as closer acceptable fits between reciprocating and stationary members which in turn results in lower bearing contact stresses.
Low pressure static sealing at the head end of the primary hydraulic structure is provided by an O-ring situated in a groove located at the interface of the cylinder bottom flange and the cylinder head reaction plate. The geometric consistancy, and therefore the sealing integrity, of this face seal application is maintained by the flange screws which prevent relative motion of the two structural members which must transmit the primary transverse loads associated with the system. Furthermore, with the exception of a very minimal squareness consideration pertaining to the high pressure sealing ring, the geometry of the face seal has no bearing on other functional interfaces, and therefore the precision of its fabrication influences no other functional fit.
Low pressure static sealing at the rod end of the system is accomplished at the cylinder bore by the exterior of the low pressure dynamic rod seal. The sectional thickness of this seal should be relatively great since the thickness influences the degree of memory which can be realized in an elastomer seal. The greater the thickness, the more the memory provided the diametrical squeeze is increased. While this squeeze increases seal drag, and therefore seal wear, this type of seal design is usually optimized at some sectional thickness as provided by the rod and cylinder bore diameters shown in the subject device. A somewhat greater seal sectional thickness could be accommodated through a snap in installation approach by undercutting the cylinder wall at the seal location.
The optimum use within the system of axial space is extremely important in providing outstanding endurance characteristics. In addition to the large bearing spacing mentioned before, the device should be able to accommodate a heavy duty rod wiper for exclusion of foreign matter and a rod seal of substantial length. The rod seal length provides stability to the seal thus enhancing its sealing capability as well as its resistance to wear. Innovations for effectively utilizing internal axial space are incorporated at both the top and bottom end of the cylinder bore. At the upper end two relatively closely spaced grooves are undercut into the cylinder wall. The uppermost groove captivates the relatively thin retainer ring which serves as the extension stop for the reciprocating members. The lower groove intercepts the diagonally running passages in the cylinder upper flange which provide communication between the discharge ends of the cylinder bore and the reservoir. This interception obviates the need to fabricate the diagonal passages all the way to the cylinder wall thus further conserving axial space commitments. Optimal use of axial space at the head end of the cylinder is provided by specific structural compaction of secondary functions at this end of the cylinder. In addition to the low profile height offered by the high pressure seal and valve seat ring, the means of captivating the return flow valve plate provide for additional axial space efficiency. Because the plate is very thin (approximately 0.010 inches thick) it can accommodate a rigid constraint without provision of any guides, stops, or auxiliary springs and utilize elastic deformation solely as the means of affording relatively free return flow of fluid. The plate flexibility also accommodates any small eccentricities between the ring seat and locating lug as well as small parallelism deviations between parts without incurring significant additional stresses. Furthermore, the return flow checking function can exhibit a small preload merely by incorporating a slight height difference between the valve seat portion of the ring seal and the central region of the reaction head. The valve plate is rigidly captivated against this region which supports most of the plate during the pressure phase of the hydraulic cycle. The captivation means can be accomplished through the use of a stand-off ring, washers, and screw respectively situated around, on, and in the central locating lug. Other similar means may be used for the valve plate captivation, but the basic configuration and approach lends itself in general to the effective use of very small components for this function. As a result, these components can be designed to fit into the space surrounded by the heads of the screws used to secure the rod to the piston. This fit is realized at full downward excursion of the rod and piston and provides one of the major contributions to axial space efficiency. The valve plate locating lug may incorporate, as shown in the illustrations, a backstop member to prevent possible excessive deformation of the valve plate should the rod and car bolster become separated during return stroke motion and the machine exhibit a momentary hang-up followed by rapid acceleration induced by the return spring.
The reservoir for the hydraulic fluid is defined by part of the outside wall of the main cylinder member and a reservoir member which encircles that part but sustains none of the system primary loads. The two members meet at a pair of equal diameter cylindrical geometries. This provides for the economical fabrication of a precision fit at these low pressure static sealing interfaces. O-rings provide fluid seals where the faces abut. The reservoir member is held in place merely by reason of its bottom abutting the cylinder head and its top abutting a retaining ring.
The passageway for the discharge of hydraulic fluid from the cylinder into the reservoir passes upwardly through the cylinder head and outwardly to a chamber about the piston rod and thence through passages into the reservoir. The fact that there is no return flow through the piston from the compensation region to the working region of the cylinder volume and the fact that all hydraulic fluid discharged from the cylinder enters the reservoir from above, while the return passageway withdraws hydraulic fluid from the bottom of the reservoir means that the heated hydraulic fluid is constantly being withdrawn and replaced with cooled hydraulic fluid from the reservoir. Furthermore, during each cycle of machine operation a full working volume, i.e., piston swept volume, is circulated through the reservoir. In usual applications of this type of device only a dunnage, or rod swept volume, of fluid is circulated through a reservoir region during each cycle. This increased cyclical circulation of hot fluid through that region where the majority of heat transfer to the environment occurs results in lower operating temperatures. The discharge passageway from the cylinder includes at least two, different orifices in a plate secured to the bottom of the piston by the same bolts that connect the piston and piston rod. The valve members of two check valves within the piston bear against the top of the orifice plate. One of these orifices is quite small and the spring of the check valve associated therewith is weak so that the check valve does little more than serve to permit air to be exhausted from the cylinder, while most of the control of the flow of hydraulic fluid from the cylinder is performed by the larger orifice. Alternately, that spring can be omitted and reliance placed on gravity to seat the ball on the seat.
By reason of features already mentioned, and the fact that the return spring for the piston is external to the cylinder, a substantial distance between the bearing areas on the piston rod and the bearing areas on the piston is achieved. Thus, there is improved performance so far as non-axial loading is concerned. The existence of the return spring outside the cylinder permits the use of a stronger spring. Thus, the contact force of the seals employed in conjunction with the piston and piston rod can be increased for better sealing and need not be compromised such as is often the case if a relatively low strength return spring is employed.
Further objects and advantages will become apparent from the following description and drawings.